1. Field of the Invention
The present invention relates in general to turbomachineries such as centrifugal and mixed flow pumps, gas blowers and compressors, and relates in particular to a turbomachinery having variable angle flow guiding device.
2. Description of the Related Art
Turbomachineries, generally referred to as pumps hereinbelow, are sometimes provided with diffusers for converting the dynamic energy of flowing fluid discharged from an impeller efficiently into a static pressure. The diffuser can be with or without vanes, but those with vanes are mostly designed simply to utilize the flow passages between the adjacent vanes as expanding flow passages.
A report entitled "Low-Solidity Cascade Diffuser" (Transaction of The Japan Society of Mechanical Engineers, Vol 45, No. 396, S54-8) described an improvement in pump performance when the pitch of the vanes is increased by making the vane cord length smaller than a value obtained by dividing the circumference length by the number of vanes. However, the vanes in this report are fixed vanes. Experiments in which vane angles are varied have been reported in "Experimental Results on a Rotatable Low Solidity Vaned Diffuser", ASME, paper 92-GT-19.
Furthermore, when the conventional centrifugal or mixed flow pump is operated at a flow rate much less than a design flow rate, flow separation occur at the impeller, diffuser and other locations in the operating system, causing a drop in the pressure rise to a value below the maximum pressure of the pump to lead to instability in the pump system (such a phenomenon as termed surge) eventually disabling a stable operation of the pumping system.
The instability phenomenon is examined in more detail in the following.
The velocity vectors of the flow discharged from the impeller can be divided into radial components and peripheral velocity components as illustrated in FIG. 1. Assuming that there is no loss in the diffuser and that the fluid is incompressible, then the quantity r.sub.2 v.theta..sub.2, which is a product of the radius at the diffuser entrance r.sub.2 and the peripheral velocity components V.theta..sub.2, is maintained to the diffuser exit according to the law of conservation of angular momentum, therefore, the peripheral velocity components V.theta..sub.3 is given by: EQU V.theta..sub.3 =V.theta..sub.2 .multidot.(r.sub.2 /r.sub.3).
where r.sub.3 is the radius at the diffuser exit. It can be seen that the velocity is reduced by the ratio of the inlet and exit radii of a diffuser.
On the other hand, the area A.sub.2 of the diffuser inlet is given by: EQU A.sub.2 =2.pi.b.sub.2 r.sub.2
where b is the width of the diffuser.
Similarly, the area A.sub.3 of the diffuser exit is given by: EQU A.sub.3 =2.pi.b.sub.3 r.sub.3
If the diffuser is a parallel-wall vaneless type diffuser, then the ratio of the areas A.sub.2 /A.sub.3 is the same as the ratio of the radii r.sub.2 /r.sub.3. Assuming that there is no loss within the diffuser and that the fluid is incompressible, the radial velocity V.sub.r3 at the diffuser exit is given by the law of conservation of mass flow as follows. EQU V.sub.r3 =V.sub.r2 .multidot.(r.sub.2 /r.sub.3)
It follows that the radial velocity component is also reduced by the ratio of the inlet/exit radii of the diffuser, and the inlet flow angle .alpha..sub.2 becomes equal to the exit flow angle .alpha..sub.3, and the flow pattern becomes an logarithmic spiral flow.
Assuming that the slip effect of the flow inside the impeller is approximately constant regardless of the flow rate, when the flow rate is progressively lowered, although the velocity component in the peripheral direction hardly changes, the radial velocity component decreases nearly proportionally to the flow rate, and the flow angle decreases.
When the flow rate is lowered even further, the flow which maintained the radial velocity component at the diffuser inlet also decreases due to the diffuser area expansion, and the radial velocity component at the diffuser exit becomes low in accordance with the law of conservation of mass flow.
Further consideration is that a boundary layer exists at the diffuser wall surface, in which both the flow velocity and the energy values are lower than those in the main flow, therefore, even if the radial velocity component is positive at the main flow, flow separation can occur within the boundary layer, and a negative velocity component is generated, and eventually develops into a large-scale reverse flow.
It is becoming clear through various investigations that the reverse flow region becomes a propagating stall accompanied by cyclic fluctuation in flow velocity and acts as a trigger to generate a large scale surge phenomenon in the entire operating system.
In the conventional pumps having a fixed diffuser, it is not possible to prevent flow separation within the boundary layer or the reverse flow caused by low flow rate through the pump. To improve on such conditions, there are several known techniques based on variable diffuser width disclosed in, for example, a U.S. Pat. No. 4,378,194; U.S. Pat. No. 3,426,964; Japanese Laid-open Patent Publication No. S58-594; and Japanese Laid-open Patent Publication No. S58-12240. In other techniques, diffuser vane angles can be varied as disclosed in, for example, Japanese Laid-open Patent Publication No. S53-113308; Japanese Laid-open Patent Publication No. S54-119111; Japanese Laid-open Patent Publication No. S54-133611; Japanese Laid-open Patent Publication No. S55-123399; Japanese Laid-open Patent Publication No. S55-125400; Japanese Laid-open Patent Publication No. S57-56699; and Japanese Laid-open Patent Publication No. H3-37397.
Although the method based on decreasing the diffuser width improve the above mentioned problem, the frictional loss at the diffuser wall increases, causing the efficiency of the diffuser to be greatly diminished. Therefore, this type of approach presents a problem that it is applicable only to a narrow range of flow rates.
Another approach based on variable angle diffuser vanes presents a problem that because the diffuser vanes are long, the diffuser vanes touch each other at some finite angle, and therefore, it is not possible to control the flow rate down to the shut-off flow rate.
The other approach disclosed in U.S. Pat. No. 3,957,392 is based on divided diffuser vanes where only an upstream portion thereof is movable, however, it is not possible to control the flow rate down to the shut-off flow rate.
Another problem presented by the variable angle diffuser vanes is that because the purpose is to optimize the performance near some design flow rate, it is not possible to control the pumping operation at or below a flow rate to cause surge. Furthermore, none of these references discloses a clear method of determining the diffuser vane angle, and therefore, they have not contributed to solving the problems of surge in a practical and useful way.
For example, a method of determining the diffuser vane angle has been discussed in a Japanese Laid-open Patent Publication No. H4-81598, but this reference also discloses only a conceptual guide to determining the vane angle near a design flow rate, and there is no clear disclosure related to a concrete method of determining a suitable vane angle for flow rates to the shut-off flow rate.
There are other methods known to prevent instability, for example, based on providing a separate bypass pipe (blow-off for blowers and compressors) so that when a low flow rate to the pump threatens instability in the operation of the pump, a bypass pipe can be opened to maintain the flow to the pump for maintaining the stable operation and reduce the flow to the equipment.
However, according to this method, it is necessary beforehand to estimate the flow rate to cause an instability in the operation of the pump, and to take a step to open a valve for the bypass pipe when this flow rate is reached. Therefore, according to this method, the entire fluid system cannot be controlled accurately unless the flow rate to cause the instability is accurately known. Also, it is necessary to know the operating characteristics of the turbomachinery correctly at various rotational speeds of the p in order to properly control the entire fluid system. Therefore, if the operation involves continuous changes in rotational speed of the pump, such a control technique is unable to keep up with the changing conditions of the pump operation.
Furthermore, even if the instability point is avoided by activating the valve on the bypass pipe, the operating conditions of the pump itself does not change, and the pump operates ineffectively, and it presents a wasteful energy consumption. Further, this type of approach requires installation of bypass pipes and valves, and the cost of the system becomes high.